Baffled thermoclines in thermodynamic generation cycle systems

ABSTRACT

Solid-state thermoclines with internal baffle structures are in used in place of heat exchangers in a closed thermodynamic cycle power generation or energy storage system, such as a closed Brayton cycle system. The baffles limit the conductive and/or radiative transfer of heat between a solid thermal medium within different zones defined by the baffle structures.

CROSS REFERENCE TO RELATED APPLICATION

This utility application is a Continuation of co-pending U.S. patent application Ser. No. 15/392,542, filed Dec. 28, 2016, the contents of which are incorporated herein by reference in their entirety, including but not limited to those portions concerning thermodynamic cycle power and/or energy storage.

BACKGROUND

In a heat engine or heat pump, a heat exchanger may be employed to transfer heat between a thermal storage medium and a working fluid for use with turbomachinery. The heat engine may be reversible, e.g., it may also be a heat pump, and the working fluid and heat exchanger may be used to transfer heat or cold to a plurality of thermal stores.

SUMMARY

In a closed thermodynamic cycle power generation or energy storage system, such as a reversible Brayton cycle system, a pressure vessel containing a solid thermal medium in a thermocline arrangement may be used as a parallel flow or counter-flow direct-contact heat exchanger in place of a fluid-to-fluid heat exchangers. The thermocline may be maintained through the use of internal baffles which segregate the solid thermal medium into zones within the pressure vessel.

Example thermocline vessels may include an insulated pressure vessel, an inlet for receiving a working fluid at non-atmospheric pressure, an outlet for dispatching the working fluid at non-atmospheric pressure, a solid thermal medium within the insulated pressure vessel and having porosity sufficient to allow the working fluid to flow through the solid thermal medium, a plurality of baffle structures defining a plurality of zones within the insulated pressure vessel, wherein the solid thermal medium is located within the plurality of zones, wherein each baffle structure is configured to limit direct transfer of heat between the solid thermal medium in different zones, one or more fluid channels configured to channel the working fluid past the baffles and in contact with the solid thermal medium.

Example systems may include a compressor, a first thermocline vessel comprising a plurality of zones of a solid thermal medium defined by baffle structures in the interior of the first thermocline vessel, wherein each baffle structure is configured to limit direct transfer of heat between the solid thermal medium in different zones, a turbine, and a working fluid circulating, in order, through (i) the compressor, (ii) the first thermocline vessel and the solid thermal medium in the interior of the first thermocline vessel, and (iii) the turbine, wherein the solid thermal medium within a first zone proximate to an inlet of the working fluid to the first thermocline vessel is at a first temperature, and wherein the solid thermal medium within a second zone proximate to an outlet of the working fluid from the first thermocline vessel is at a second temperature higher than the first temperature.

Example energy generation systems may include a generator configured to receive mechanical energy and generate electrical energy a working fluid, a compressor configured to compress the working fluid, a turbine configured to convert expansion of the working fluid within the turbine into mechanical energy, wherein the turbine is mechanically coupled to the compressor and is configured to transmit a portion of the mechanical energy through the mechanical coupling to drive the compressor, a recuperative heat exchanger configured to thermally contact working fluid exiting the compressor with working fluid exiting the turbine; a cooling tower configured to eject to the atmosphere heat carried by the working fluid, a hot-side thermocline vessel comprising a plurality of zones of a first solid thermal medium defined by baffle structures in the interior of the first thermocline vessel, wherein each baffle structure is configured to limit direct transfer of heat between the first solid thermal medium in different zones, and a cold-side thermocline vessel comprising a plurality of zones of a second solid thermal medium defined by baffle structures in the interior of the first thermocline vessel, wherein each baffle structure is configured to limit direct transfer of heat between the second solid thermal medium in different zones, wherein the working fluid circulates, in order, through (i) the compressor, (ii) the recuperative heat exchanger, (iii) the hot-side thermocline vessel and the first solid thermal medium in the interior of the hot-side thermocline vessel, (iv) the turbine, (v) the recuperative heat exchanger, (vi) the cooling tower, and (vii) the cold-side thermocline vessel and the second solid thermal medium in the interior of the cold-side thermocline vessel.

These as well as other aspects, advantages, and alternatives, will become apparent to those of ordinary skill in the art by reading the following detailed description, with reference where appropriate to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic flow diagram of working fluid and heat storage media of a thermal system in a charge/heat pump mode.

FIG. 2 is a schematic flow diagram of working fluid and heat storage media of a thermal system in a discharge/heat engine mode.

FIG. 3A is a schematic pressure and temperature diagram of the working fluid as it undergoes the charge cycle in FIG. 1 .

FIG. 3B is a schematic pressure and temperature diagram of the working fluid as it undergoes the discharge cycle in FIG. 2 .

FIG. 4 is a schematic flow diagram of working fluid and heat storage media of a thermal system with a gas-gas heat exchanger for the working fluid in a charge/heat pump mode.

FIG. 5 is a schematic flow diagram of working fluid and heat storage media of a thermal system with a gas-gas heat exchanger for the working fluid in a discharge/heat engine mode.

FIG. 6A is a schematic pressure and temperature diagram of the working fluid as it undergoes the charge cycle in FIG. 4 .

FIG. 6B is a schematic pressure and temperature diagram of the working fluid as it undergoes the discharge cycle in FIG. 5 .

FIG. 7 illustrates a schematic flow diagram according to an example embodiment.

FIG. 8 illustrates a schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment.

FIG. 9 illustrates a schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment.

FIG. 10 illustrates a schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment.

FIG. 11 illustrates a schematic arrangement, in cut-away view, of a baffled thermocline vessel with a runner, according to an example embodiment.

FIG. 12 illustrates an example embodiment of a baffle structures inside a thermocline pressure vessel.

FIG. 13 illustrates an example embodiment of a baffle structures inside a thermocline pressure vessel.

FIG. 14 illustrates an example embodiment of a baffle structures inside a thermocline pressure vessel.

FIG. 15 illustrates an example embodiment of a baffle structures inside a thermocline pressure vessel.

DETAILED DESCRIPTION

Example methods and systems are described herein. It should be understood that the words “example” and/or “exemplary” are used herein to mean “serving as an example, instance, or illustration.” Any embodiment or feature described herein as being an “example” or “exemplary” is not necessarily to be construed as preferred or advantageous over other embodiments or features. The example embodiments described herein are not meant to be limiting. It will be readily understood that certain aspects of the disclosed systems and methods can be arranged and combined in a wide variety of different configurations, all of which are contemplated herein.

I. Overview

Reversible heat engines may use one or more solid thermal storage mediums to transfer heat to or from a working fluid. A solid thermal storage medium may be contained within a thermocline vessel where the solid thermal storage medium exists in a thermocline state with a temperature gradient across the medium. Disclosed herein are baffled thermocline vessels that may be used to help beneficially maintain that temperature gradient. An example reversible closed heat engine in which baffled thermocline vessels may be implemented is a Brayton engine system. A Brayton engine system may use a generator/motor connected to a turbine and a compressor, where the turbomachinery acts on a working fluid circulating in the system. Non-comprehensive examples of working fluids include air, argon, carbon dioxide, or gaseous mixtures. A Brayton system may have a hot side and a cold side. Each side may include a heat exchanger vessel containing a solid thermal medium The solid thermal medium may take many forms, including but not limited to, dirt, rock, gravel, sand, clay, metal, metal oxide, refractory material, refractory metal, ceramic, cermet, alumina, silica, magnesia, zirconia, silicon carbide, titanium carbide, tantalum carbide, chromium carbide, niobium carbide, zirconium carbide, molybdenum disilicide, calcium oxide, chromite, dolomite, magnesite, quartzite, aluminum silicate, tungsten, molybdenum, niobium, tantalum, rhenium, beryllium, and combinations thereof. The solid thermal medium for use in cold systems may further include water ice, and/or other solid forms of common room temperature liquids. Preferably, the solid medium is structurally stable at high or low temperature, of uniform shape and/or size, and shaped such that a bolus of the solid medium includes gaps to allow a working fluid to flow through the bolus. For example, for refractory materials it may be preferable to utilize larges slabs, stackable bricks, platonic solids, spheres, cylinders, or other shapes that can be stacked and/or arranged to allow gaps between individual units of the the solid medium. For metal, metal oxides, or ceramics it may be preferable to use those shapes or fabrics or meshes that consist entirely or partially of the metal, metal oxide, or ceramic, where the fabric or mesh has a porosity sufficient to allow passage of a working fluid through the solid medium.

The hot-side solid thermal medium may reach temperatures over 600° C. and, if the heat exchanger vessel operates as direct contact between the working fluid and the hot-side solid thermal medium, the pressure may be over over 100 bars. Similarly, cold-side thermal medium can go below −70° C. and be at or near vacuum state in the heat exchanger.

It may be desirable to configure a Brayton cycle heat exchanger vessel as a counter-flow heat exchanger to maximize efficiency of the thermal cycle. Preferably, this may be implemented via a thermocline, or temperature gradient within the heat exchanger vessel, wherein the heat exchanger vessel may include a pressure vessel. For a hot-side heat exchanger vessel in a discharging Brayton cycle, the entering working fluid preferably contacts the coldest solid thermal medium at the entrance to the heat exchanger vessel and contacts the hottest solid thermal medium at the exit to the heat exchanger vessel. For a cold-side heat exchanger in a discharging Brayton cycle, the entering working fluid preferably contacts the hottest solid thermal medium at the entrance to the heat exchanger and contacts the coldest solid thermal medium at the exit to the heat exchanger. The contact between the working fluid and the solid thermal medium may be direct contact or indirect thermal contact depending on the configuration. In a charging Brayton cycle, the contacting order would preferably be reversed.

Disadvantageously, the solid thermal medium in a thermocline arrangement within a heat exchanger vessel may conduct, convect, and/or radiate heat from hot portions of the solid thermal medium to cold portions until equilibrium is reached throughout the solid thermal medium. Non-beneficial direct transfer of heat through the solid thermal medium may include conductive heat transfer from a portion of the solid thermal medium in direct contact with another portion of the solid thermal medium, or in contact through a thermal transfer medium such as an uninsulated and/or conductive surface of the heat exchanger vessel. Direct transfer of heat may also include radiative (or emissive) heat transfer from a portion of solid thermal medium to another portion of solid thermal medium, where the two portions are not in direct contact. For purposes herein, direct transfer of heat does not include heat transfer by means of a working fluid passing over solid thermal medium, where the working fluid carries heat from one portion of the solid thermal medium to another portion. Prior to thermal equilibrium, conductive and/or radiative heat transfer through or among the solid thermal medium will reduce the maximum temperature difference across the thermocline and potentially reduce overall thermal efficiency of a thermodynamic (e.g. Brayton) cycle employing a direct contact thermocline heat exchanger.

To mitigate direct heat transfer within the solid thermal storage medium, one or more baffle structures may be included in the heat exchanger vessel to create zones within the heat exchanger vessel. Each zone may contain a portion of the total quantity of the solid thermal medium in the heat exchanger vessel and the baffle structures may limit direct transfer of heat between the solid thermal medium in one zone to the solid thermal medium in a different zone.

The baffle structures may take forms configured to limit direct transfer of heat between the solid thermal medium in different zones, while still allowing a working fluid to reach the the solid thermal medium while flowing through the heat exchanger. For example, a baffle structure may be an insulated wall extending partially across the heat exchanger, leaving a space for working fluid flow between an end of the insulated wall and an interior wall of the pressure vessel. As another example, a baffle structure may be an insulated wall extending completely across the heat exchanger. The wall may include fluid channels through the wall that allow a working fluid to flow through the wall. The working fluid may flow throughout the heat exchanger or may be constrained to runners that traverse the heat exchanger and are in thermal contact with the solid thermal medium.

The fluid channels may be sized or located to limit or prevent the movement of the solid thermal medium through the fluid channels. Alternatively or additionally, the fluid channels may be sized, located, and/or shaped to limit the effect of radiative heat transfer. For example, a fluid channel may be angled or convoluted in form such that there is no, or limited, line-of-sight between the solid thermal medium on one side of the fluid channel and the solid thermal medium on the other side of the fluid channel. Similarly, another baffle structure may be a perforated material extending across the heat exchanger. The perforated material may be physically separated from the the solid thermal medium to prevent or limit conductive heat transfer and the perforations may act similarly to the wall-based fluid channels described above. Other baffle structures may include insulated chambers within the heat exchanger that each contain the solid thermal medium and a plurality of fluid channels that allow a working fluid to enter the chamber, contact the solid thermal medium directly or through runners, and exit the chamber.

II. Illustrative Reversible Heat Engine

Systems and devices in which example embodiments may be implemented will now be described in greater detail. However, an example system may also be implemented in or take the form of other devices, without departing from the scope of the invention.

An aspect of the disclosure relates to thermal systems operating on thermal storage cycles. In some examples, the cycles allow electricity to be stored as heat (e.g., in the form of a temperature differential) and then converted back to mechanical work and ultimately electricity through the use of at least two pieces of turbomachinery (a compressor and a turbine), and a generator. The compressor consumes work and raises the temperature and pressure of a working fluid (WF). The turbine produces work and lowers the temperature and pressure of the working fluid. In example systems, more than one compressor and/or more than one turbine may be used. The compressors may be arranged in series or in parallel. The turbines may be arranged in series or in parallel.

FIGS. 1 and 2 are schematic flow diagrams of working fluid and heat storage medium of an example thermal system in a charge/heat pump mode and in a discharge/heat engine mode, respectively. The system may be idealized for simplicity of explanation so that there are no losses (i.e., entropy generation) in either the turbomachinery or heat exchangers. The system can include a working fluid (e.g., argon gas) flowing in a closed cycle between a compressor 1, a hot side heat exchanger 2, a turbine 3 and a cold side heat exchanger 4. Fluid flow paths/directions for the working fluid (e.g., a gas), a hot side thermal storage (HTS) medium 21 (e.g., a low viscosity liquid or a solid medium) and a cold side thermal storage (CTS) medium 22 (e.g., a low viscosity liquid or a solid medium, which may be different from the HTS medium) are indicated by arrows. The heat exchangers 2 and 4 exchangers may incorporate, for example, conventional liquid-to-gas exchange for liquid thermal storage media (e.g., tube-and-shell exchangers or plate exchanger) and solid-to-gas exchange (e.g., direct contact) for the solid thermal medium and may require pumping and/or conveyance mechanisms for the media.

FIGS. 3A and 3B are schematic pressure and temperature diagrams of the working fluid as it undergoes the charge cycles in FIGS. 1 and 2 , respectively, once again simplified in the approximation of no entropy generation. Normalized pressure is shown on the y-axes and temperature is shown on the x-axes. The direction of processes taking place during the cycles is indicated with arrows, and the individual processes taking place in the compressor 1, the hot side CFX 2, the turbine 3 and the cold side CFX 4 are indicated on the diagram with their respective circled numerals.

The heat exchangers 2 and 4 can be configured as counter-flow heat exchangers (CFXs), where the working fluid flows in one direction and the substance it is exchanging heat with is flowing or moving or has a temperature gradient in the opposite direction. In an ideal counter-flow heat exchanger with correctly matched flows (i.e., balanced capacities or capacity flow rates or thermocline gradient), the temperatures of the working fluid and thermal storage medium flip (i.e., the counter-flow heat exchanger can have unity effectiveness).

The counter-flow heat exchangers 2 and 4 can be designed and/or operated to reduce entropy generation in the heat exchangers to negligible levels compared to entropy generation associated with other system components and/or processes (e.g., compressor and/or turbine entropy generation). In some cases, the system may be operated such that entropy generation in the system is minimized. For example, the system may be operated such that entropy generation associated with heat storage units is minimized. In some cases, a temperature difference between fluid or solid elements exchanging heat can be controlled during operation such that entropy generation in hot side and cold side heat storage units is minimized. In some instances, the entropy generated in the hot side and cold side heat storage units is negligible when compared to the entropy generated by the compressor, the turbine, or both the compressor and the turbine. In some instances, entropy generation associated with heat transfer in the heat exchangers 2 and 4 and/or entropy generation associated with operation of the hot side storage unit, the cold side storage unit or both the hot side and cold side storage units can be less than about 50%, less than about 25%, less than about 20%, less than about 15%, less than about 10%, less than about 5%, less than about 4%, less than about 3%, less than about 2%, or less than about 1% of the total entropy generated within the system (e.g., entropy generated by the compressor 1, the hot side heat exchanger 2, the turbine 3, the cold side heat exchanger 4 and/or other components described herein, such as, for example, a recuperator). For example, entropy generation can be reduced or minimized if the two substances exchanging heat do so at a local temperature differential ΔT→0 (i.e., when the temperature difference between any two fluid or solid media elements that are in close thermal contact in the heat exchanger is small). In some examples, the temperature differential ΔT between any two fluid or solid media elements that are in close thermal contact may be less than about 300 Kelvin (K), less than about 200 K, less than about 100 K, less than about 75 K, less than about 50 K, less than about 40 K, less than about 30 K, less than about 20 K, less than about 10 K, less than about 5 K, less than about 3 K, less than about 2 K, or less than about 1 K. In another example, entropy generation associated with pressure drop can be reduced or minimized by suitable design. In some examples, the heat exchange process can take place at a constant or near-constant pressure. Alternatively, a non-negligible pressure drop may be experienced by the working fluid and/or one or more thermal storage media during passage through a heat exchanger. Pressure drop in heat exchangers may be controlled (e.g., reduced or minimized) through suitable heat exchanger design. In some examples, the pressure drop across each heat exchanger may be less than about 20% of inlet pressure, less than about 10% of inlet pressure, less than about 5% of inlet pressure, less than about 3% of inlet pressure, less than about 2% of inlet pressure, less than about 1% of inlet pressure, less than about 0.5% of inlet pressure, less than about 0.25% of inlet pressure, or less than about 0.1% of inlet pressure.

Upon entering the heat exchanger 2, the temperature of the working fluid can either increase (taking heat from the HTS medium 21, corresponding to the discharge mode in FIGS. 2 and 3B) or decrease (giving heat to the HTS medium 21, corresponding to the charge mode in FIGS. 1 and 3A), depending on the temperature of the HTS medium in the heat exchanger relative to the temperature of the working fluid. Similarly, upon entering the heat exchanger 4, the temperature of the working fluid can either increase (taking heat from the CTS medium 22, corresponding to the charge mode in FIGS. 1 and 3A) or decrease (giving heat to the CTS medium 22, corresponding to the discharge mode in FIGS. 2 and 3B), depending on the temperature of the CTS medium in the heat exchanger relative to the temperature of the working fluid.

As described in more detail with reference to the charge mode in FIGS. 1 and 3A, the heat addition process in the cold side CFX 4 can take place over a different range of temperatures than the heat removal process in the hot side CFX 2. Similarly, in the discharge mode in FIGS. 2 and 3B, the heat rejection process in the cold side CFX 4 can take place over a different range of temperatures than the heat addition process in the hot side CFX 2. At least a portion of the temperature ranges of the hot side and cold side heat exchange processes may overlap during charge, during discharge, or during both charge and discharge.

As used herein, the temperatures T₀, T₁, T₀ ⁺ and T₁ ⁺ are so named because T₀ ^(t), T₁ ⁺ are the temperatures achieved at the exit of a compressor with a given compression ratio r, adiabatic efficiency η_(c) and inlet temperatures of T₀, T₁ respectively. The examples in FIGS. 1, 2, 3A and 3B can be idealized examples where η_(c)=1 and where adiabatic efficiency of the turbine η_(t) also has the value η_(t)=1.

With reference to the charge mode shown in FIGS. 1 and 3A, the working fluid can enter the compressor 1 at position 30 at a pressure P and a temperature T (e.g., at T₁, P₂). As the working fluid passes through the compressor, work W₁ is consumed by the compressor to increase the pressure and temperature of the working fluid (e.g., to T₁ ⁺, P₁), as indicated by PT and T↑ at position 31. In the charge mode, the temperature T₁ ⁺ of the working fluid exiting the compressor and entering the hot side CFX 2 at position 31 is higher than the temperature of the HTS medium 21 entering the hot side CFX 2 at position 32 from a second hot side thermal storage tank 7 at a temperature T₀ ⁺ (i.e., T₀ ⁺<T₁ ⁺). As these working fluid and thermal medium pass in thermal contact with each other in the heat exchanger, the working fluid's temperature decreases as it moves from position 31 to position 34, giving off heat Q₁ to the HTS medium, while the temperature of the HTS medium in turn increases as it moves from position 32 to position 33, absorbing heat Q1 from the working fluid. In an example, the working fluid exits the hot side CFX 2 at position 34 at the temperature T₀ ⁺ and the HTS medium exits the hot side CFX 2 at position 33 into a first hot side thermal storage tank 6 at the temperature T₁ ⁺. The heat exchange process can take place at a constant or near-constant pressure such that the working fluid exits the hot side CFX 2 at position 34 at a lower temperature but same pressure P₁, as indicated by P and T↓, at position 34. Similarly, the temperature of the HTS medium 21 increases in the hot side CFX 2, while its pressure can remain constant or near-constant.

Upon exiting the hot side CFX 2 at position 34 (e.g., at T₀ ⁺, P₁), the working fluid undergoes expansion in the turbine 3 before exiting the turbine at position 35. During the expansion, the pressure and temperature of the working fluid decrease (e.g., to T₀, P₂), as indicated by P↓, and T↓ at position 35. The magnitude of work W₂ generated by the turbine depends on the enthalpy of the working fluid entering the turbine and the degree of expansion. In the charge mode, heat is removed from the working fluid between positions 31 and 34 (in the hot side CFX 2) and the working fluid is expanded back to the pressure at which it initially entered the compressor at position 30 (e.g., P₂). The compression ratio (e.g., P₁/P₂) in the compressor 1 being equal to the expansion ratio in the turbine 3, and the enthalpy of the gas entering the turbine being lower than the enthalpy of the gas exiting the compressor, the work W₂ generated by the turbine 3 is smaller than the work W₁ consumed by the compressor 1 (i.e., W₂<W₁).

Because heat was taken out of the working fluid in the hot side CFX 2, the temperature T₀ at which the working fluid exits the turbine at position 35 is lower than the temperature T₁ at which the working fluid initially entered the compressor at position 30. T₀ close the cycle (i.e., to return the pressure and temperature of the working fluid to their initial values T₁, P₂ at position 30), heat Q₂ is added to the working fluid from the CTS medium 22 in the cold side CFX 4 between positions 35 and 30 (i.e., between the turbine 3 and the compressor 1). In an example, the CTS medium 22 enters the cold side CFX 4 at position 36 from a first cold side thermal storage tank 8 at the temperature T₁ and exits the cold side CFX 4 at position 37 into a second cold side thermal storage tank 9 at the temperature T₀, while the working fluid enters the cold side CFX 4 at position 35 at the temperature T₀ and exits the cold side CFX 4 at position 30 at the temperature T₁. Again, the heat exchange process can take place at a constant or near-constant pressure such that the working fluid exits the cold side CFX 2 at position 30 at a higher temperature but same pressure P₂, as indicated by P and TT at position 30. Similarly, the temperature of the CTS medium 22 decreases in the cold side CFX 2, while its pressure can remain constant or near-constant.

During charge, the heat Q₂ is removed from the CTS medium and the heat Q₁ is added to the HTS medium, wherein Q₁>Q₂. A net amount of work (W₁−W₂) is consumed, since the work W₁ used by the compressor is greater than the work W₂ generated by the turbine. A device that consumes work while moving heat from a cold body or thermal storage medium to a hot body or thermal storage medium may be considered a heat pump; thus, the thermal system in the charge mode may operate as a heat pump.

In an example, the discharge mode shown in FIGS. 2 and 3B can differ from the charge mode shown in FIGS. 1 and 3A in the temperatures of the thermal storage media being introduced into the heat exchangers. The temperature at which the HTS medium enters the hot side CFX 2 at position 32 is T₁ ⁺ instead of T₀ ⁺, and the temperature of the CTS medium entering the cold side CFX 4 at position 36 is T₀ instead of T₁. During discharge, the working fluid enters the compressor at position 30 at T₀ and P₂, exits the compressor at position 31 at T₀ ⁺<T₁ ⁺ and P₁, absorbs heat from the HTS medium in the hot side CFX 2, enters the turbine 3 at position 34 at T₁ ⁺ and P₁, exits the turbine at position 35 at T₁>T₀ and P₂, and finally rejects heat to the CTS medium in the cold side CFX 4, returning to its initial state at position 30 at T₀ and P₂.

The HTS medium at temperature T₁ ⁺ can be stored in a first hot side thermal storage tank 6, the HTS medium at temperature T₀ ⁺ can be stored in a second hot side thermal storage tank 7, the CTS medium at temperature T₁ can be stored in a first cold side thermal storage tank 8, and the CTS medium at temperature T₀ can be stored in a second cold side thermal storage tank 9 during both charge and discharge modes. In one implementation, the inlet temperature of the HTS medium at position 32 can be switched between T₁ ⁺ and T₀ ⁺ by switching between tanks 6 and 7, respectively. Similarly, the inlet temperature of the CTS medium at position 36 can be switched between T₁ and T₀ by switching between tanks 8 and 9, respectively. Switching between tanks can be achieved by including a valve or a system of valves, or a conveyance system or a group of conveyance systems, for switching connections between the hot side heat exchanger 2 and the hot side tanks 6 and 7, and/or between the cold side heat exchanger 4 and the cold side tanks 8 and 9 as needed for the charge and discharge modes. In some implementations, connections may be switched on the working fluid side instead, while the connections of storage tanks 6, 7, 8 and 9 to the heat exchangers 2 and 4 remain static. In some examples, flow paths and connections to the heat exchangers may depend on the design (e.g., shell-and-tube or direct-contact) of each heat exchanger. In some implementations, one or more valves or conveyance systems can be used to switch the direction of both the working fluid and the heat storage media through the counter-flow heat exchanger on charge and discharge. Such configurations may be used, for example, due to high thermal storage capacities of the heat exchanger component, to decrease or eliminate temperature transients, or a combination thereof. In some implementations, one or more valves or conveyance systems can be used to switch the direction of only the working fluid, while the direction of the HTS or CTS can be changed by changing the direction of pumping or conveyance, thereby maintaining the counter-flow configuration. In some implementations, different valve configurations or conveyance systems may be used for the HTS and the CTS. Further, any combination of the valve or conveyance configurations herein may be used. For example, the system may be configured to operate using different valve or conveyance configurations in different situations (e.g., depending on system operating conditions).

In the discharge mode shown in FIGS. 2 and 3B, the working fluid can enter the compressor 1 at position 30 at a pressure P and a temperature T (e.g., at T₀, P₂). As the working fluid passes through the compressor, work W₁ is consumed by the compressor to increase the pressure and temperature of the working fluid (e.g., to T₀ ⁺, P₁), as indicated by PT and TT at position 31. In the discharge mode, the temperature T₀ ⁺ of the working fluid exiting the compressor and entering the hot side CFX 2 at position 31 is lower than the temperature of the HTS medium 21 entering the hot side CFX 2 at position 32 from a first hot side thermal storage tank 6 at a temperature T₁ ⁺ (i.e., T₀ ⁺<T₁ ⁺). As these two fluids pass in thermal contact with each other in the heat exchanger, the working fluid's temperature increases as it moves from position 31 position 34, absorbing heat Q₁ from the HTS medium, while the temperature of the HTS medium in turn decreases as it moves from position 32 to position 33, giving off heat Q₁ to the working fluid. In an example, the working fluid exits the hot side CFX 2 at position 34 at the temperature T₁ ⁺ and the HTS medium exits the hot side CFX 2 at position 33 into the second hot side thermal storage tank 7 at the temperature T₀ ⁺. The heat exchange process can take place at a constant or near-constant pressure such that the working fluid exits the hot side CFX 2 at position 34 at a higher temperature but same pressure P₁, as indicated by P and TT at position 34. Similarly, the temperature of the HTS medium 21 decreases in the hot side CFX 2, while its pressure can remain constant or near-constant.

Upon exiting the hot side CFX 2 at position 34 (e.g., at T₁ ⁺, P₁), the working fluid undergoes expansion in the turbine 3 before exiting the turbine at position 35. During the expansion, the pressure and temperature of the working fluid decrease (e.g., to T₁, P₂), as indicated by P_(↓) and T_(↓) at position 35. The magnitude of work W₂ generated by the turbine depends on the enthalpy of the working fluid entering the turbine and the degree of expansion. In the discharge mode, heat is added to the working fluid between positions 31 and 34 (in the hot side CFX 2) and the working fluid is expanded back to the pressure at which it initially entered the compressor at position 30 (e.g., P₂). The compression ratio (e.g., P₁/P₂) in the compressor 1 being equal to the expansion ratio in the turbine 3, and the enthalpy of the gas entering the turbine being higher than the enthalpy of the gas exiting the compressor, the work W₂ generated by the turbine 3 is greater than the work W₁ consumed by the compressor 1 (i.e., W₂>W₁).

Because heat was added to the working fluid in the hot side CFX 2, the temperature T₁ at which the working fluid exits the turbine at position 35 is higher than the temperature T₀ at which the working fluid initially entered the compressor at position 30. T₀ close the cycle (i.e., to return the pressure and temperature of the working fluid to their initial values T₀, P₂ at position 30), heat Q₂ is rejected by the working fluid to the CTS medium 22 in the cold side CFX 4 between positions 35 and 30 (i.e., between the turbine 3 and the compressor 1). The CTS medium 22 enters the cold side CFX 4 at position 36 from a second cold side thermal storage tank 9 at the temperature T₀ and exits the cold side CFX 4 at position 37 into a first cold side thermal storage tank 8 at the temperature T₁, while the working fluid enters the cold side CFX 4 at position 35 at the temperature T₁ and exits the cold side CFX 4 at position 30 at the temperature T₀. Again, the heat exchange process can take place at a constant or near-constant pressure such that the working fluid exits the cold side CFX 2 at position 30 at a higher temperature but same pressure P₂, as indicated by P and T↓ at position 30. Similarly, the temperature of the CTS medium 22 increases in the cold side CFX 2, while its pressure can remain constant or near-constant.

During discharge, the heat Q₂ is added to the CTS medium and the heat Q₁ is removed from the HTS medium, wherein Q₁>Q₂. A net amount of work (W₂−W₁) is generated, since the work W₁ used by the compressor is smaller than the work W₂ generated by the turbine. A device that generates work while moving heat from a hot body or thermal storage medium to a cold body or thermal storage medium is a heat engine; thus, the thermal system in the discharge mode operates as a heat engine.

Another aspect of the disclosure is directed to thermal systems with regeneration/recuperation. In some situations, the terms regeneration and recuperation can be used interchangeably, although they may have different meanings. As used herein, the terms “recuperation” and “recuperator” generally refer to the presence of one or more additional heat exchangers where the working fluid exchanges heat with itself during different segments of a thermodynamic cycle through continuous heat exchange without intermediate thermal storage. As used herein, the terms “regeneration” and “regenerator” may be used to describe the same configuration as the terms “recuperation” and “recuperator.” The roundtrip efficiency of thermal systems may be substantially improved if the allowable temperature ranges of the storage materials can be extended. In some implementations, this may be accomplished by choosing a material or medium on the cold side that can go to temperatures below 273 K (0° C.). For example, a CTS medium (e.g., hexane) with a low temperature limit of approximately T₀=179 K (−94° C.) may be used in a system with a molten salt or solid HTS medium. However, T₀ ⁺ (i.e., the lowest temperature of the working fluid in the hot side heat exchanger) at some (e.g., modest) compression ratios may be below the freezing point of the molten salt, making the molten salt unviable as the HTS medium. In some implementations, this can be resolved by including a working fluid to working fluid (e.g., gas-gas) heat exchanger (also “recuperator” or “regenerator” herein) in the cycle.

FIG. 4 is a schematic flow diagram of working fluid and heat storage media of a thermal system in a charge/heat pump mode with a gas-gas heat exchanger 5 for the working fluid. The use of the gas-gas heat exchanger can enable use of colder heat storage medium on the cold side of the system. As examples, the working fluid can be air, argon, or a mixture of primarily argon mixed with another gas such as helium. For example, the working fluid may comprise at least about 50% argon, at least about 60% argon, at least about 70% argon, at least about 80% argon, at least about 90% argon, or about 100% argon, with balance helium.

FIG. 6A shows a heat storage charge cycle for the storage system in FIG. 4 with a cold side storage medium (e.g., liquid hexane or heptane or solid thermal medium) capable of going down to approximately to 179 K (−94° C.) and a molten salt or a solid thermal medium as the hot side storage, and η_(c)=0.9 and η_(t)=0.95. In some cases, the system can include more than four heat storage tanks.

In one implementation, during charge in FIGS. 4 and 6A, the working fluid enters the compressor at T₁ and P₂, exits the compressor at T₁ ⁺ and P₁, rejects heat Q₁, to the HTS medium 21 in the hot side CFX 2, exiting the hot side CFX 2 at T₁ and P₁, rejects heat Q_(recup) (also “Q_(regen)” herein, as shown, for example, in the accompanying drawings) to the cold (low pressure) side working fluid in the heat exchanger or recuperator 5, exits the recuperator 5 at T₀ ⁺ and P₁, rejects heat to the environment (or other heat sink) in section 38 (e.g., a radiator), enters the turbine 3 at T₀ ⁺ and P₁, exits the turbine at T₀ and P₂, absorbs heat Q₂ from the CTS medium 22 in the cold side CFX 4, exiting the cold side CFX 4 at T₀ ⁺ and P₂, absorbs heat Q_(recup) from the hot (high pressure) side working fluid in the heat exchanger or recuperator 5, and finally exits the recuperator 5 at T₁ and P₂, returning to its initial state before entering the compressor.

FIG. 5 is a schematic flow diagram of working fluid and heat storage media of the thermal system in FIG. 4 in a discharge/heat engine mode. Again, the use of the gas-gas heat exchanger can enable use of colder heat storage fluid or a solid medium (CTS) and/or colder working fluid on the cold side of the system.

FIG. 6B shows a heat storage discharge cycle for the storage system for the storage system in FIG. 5 with a cold side storage medium (e.g., liquid hexane or solid thermal storage medium) capable of going down to 179 K (−94° C.) and a molten salt or a solid thermal storage medium as the hot side storage, and ηc=0.9 and ηt=0.95.

During discharge in FIGS. 5 and 6B, the working fluid enters the compressor at T₀ and P₂, exits the compressor at T₀ ⁺ and P₁, absorbs heat Q_(recup) from the cold (low pressure) side working fluid in the heat exchanger or recuperator 5, exits the recuperator 5 at T₁ and P₁, absorbs heat Q₁ from the HTS medium 21 in the hot side CFX 2, exiting the hot side CFX 2 at T₁ ⁺ and P₁, enters the turbine 3 at {tilde over (T)}₁ ⁺ and P₁, exits the turbine at {tilde over (T)}₁ and P₂, rejects heat to the environment (or other heat sink) in section 39 (e.g., a radiator), rejects heat Q_(recup) to the hot (high pressure) side working fluid in the heat exchanger or recuperator 5, enters the cold side CFX 4 at T₀ ⁺ and P₂, rejects heat Q₂ to the CTS medium 22 in the cold side CFX 4, and finally exits the cold side CFX 4 at T₀ and P₂, returning to its initial state before entering the compressor.

In some examples, recuperation may enable the compression ratio to be reduced. In some cases, reducing the compression ratio may result in reduced compressor and turbine losses. In some cases, the compression ratio may be at least about 1.2, at least about 1.5, at least about 2, at least about 2.5, at least about 3, at least about 3.5, at least about 4, at least about 4.5, at least about 5, at least about 6, at least about 8, at least about 10, at least about 15, at least about 20, at least about 30, or more.

In some cases, T₀ may be at least about 30 K, at least about 50 K, at least about 80 K, at least about 100 K, at least about 120 K, at least about 140 K, at least about 160 K, at least about 180 K, at least about 200 K, at least about 220 K, at least about 240 K, at least about 260 K, or at least about 280 K. In some cases, T₀ ⁺ may be at least about 220 K, at least about 240 K, at least about 260 K, at least about 280 K, at least about 300 K, at least about 320 K, at least about 340 K, at least about 360 K, at least about 380 K, at least about 400 K, or more. In some cases, the temperatures T₀ and T₀ ⁺ can be constrained by the ability to reject excess heat to the environment at ambient temperature due to inefficiencies in components such as turbomachinery. In some cases, the temperatures T₀ and T₀ ⁺ can be constrained by the operating temperatures of the CTS (e.g., a phase transition temperature). In some cases, the temperatures T₀ and T₀ ⁺ can be constrained by the compression ratio being used. Any description of the temperatures T₀ and/or T₀ ⁺ herein may apply to any system or method of the disclosure.

In some cases, T₁ may be at least about 350K, at least about 400 K, at least about 440 K, at least about 480 K, at least about 520 K, at least about 560 K, at least about 600 K, at least about 640 K, at least about 680 K, at least about 720 K, at least about 760 K, at least about 800 K, at least about 840 K, at least about 880 K, at least about 920 K, at least about 960 K, at least about 1000 K, at least about 1100 K, at least about 1200 K, at least about 1300 K, at least about 1400 K, or more. In some cases, T₁ ⁺ may be at least about 480 K, at least about 520 K, at least about 560 K, at least about 600 K, at least about 640 K, at least about 680 K, at least about 720 K, at least about 760 K, at least about 800 K, at least about 840 K, at least about 880 K, at least about 920 K, at least about 960 K, at least about 1000 K, at least about 1100 K, at least about 1200 K, at least about 1300 K, at least about 1400 K, at least about 1500 K, at least about 1600 K, at least about 1700 K, or more. In some cases, the temperatures T₁ and T₁ ⁺ can be constrained by the operating temperatures of the HTS. In some cases, the temperatures T₁ and T₁ ⁺ can be constrained by the thermal limits of the metals and materials being used in the system. For example, a conventional solar salt can have a recommended temperature range of approximately 560-840 K. Various system improvements, such as, for example, increased round-trip efficiency, increased power and increased storage capacity may be realized as available materials, metallurgy and storage materials improve over time and enable different temperature ranges to be achieved. Any description of the temperatures T₁ and/or T₁ ⁺ herein may apply to any system or method of the disclosure.

In some cases, the round-trip efficiency ii store (e.g., electricity storage efficiency) with and/or without recuperation can be at least about 5%, at least about 10%, at least about 15%, at least about 20%, at least about 25%, at least about 30%, at least about 35%, at least about 40%, at least about 45%, at least about 50%, at least about 55%, at least about 60%, at least about 65%, at least about 70%, at least about 75%, at least about 80%, at least about 85%, at least about 90%, or at least about 95%.

In some implementations, at least a portion of heat transfer in the system (e.g., heat transfer to and from the working fluid) during a charge and/or discharge cycle includes heat transfer with the environment (e.g., heat transfer in sections 38 and 39). The remainder of the heat transfer in the system can occur through thermal communication with thermal storage media (e.g., thermal storage media 21 and 22), through heat transfer in the recuperator 5 and/or through various heat transfer processes within system boundaries (i.e., not with the surrounding environment). In some examples, the environment may refer to gaseous or liquid reservoirs surrounding the system (e.g., air, water), any system or media capable of exchanging thermal energy with the system (e.g., another thermodynamic cycle or system, heating/cooling systems, etc.), or any combination thereof. In some examples, heat transferred through thermal communication with the heat storage media can be at least about 25%, at least about 50%, at least about 60%, at least about 70%, at least about 80%, or at least about 90% of all heat transferred in the system. In some examples, heat transferred through heat transfer in the recuperator can be at least about 5%, at least about 10%, at least about 15%, at least about 20%, at least about 25%, at least about 50%, or at least about 75% of all heat transferred in the system. In some examples, heat transferred through thermal communication with the heat storage media and through heat transfer in the recuperator can be at least about 25%, at least about 50%, at least about 60%, at least about 70%, at least about 80%, at least about 90%, or even about 100% of all heat transferred in the system. In some examples, heat transferred through heat transfer with the environment can be less than about 5%, less than about 10%, less than about 15%, less than about 20%, less than about 30%, less than about 40%, less than about 50%, less than about 60%, less than about 70%, less than about 80%, less than about 90%, less than about 100%, or even 100% of all heat transferred in the system. In some implementations, all heat transfer in the system may be with the thermal storage media (e.g., the CTS and HTS media), and only the thermal storage media may conduct heat transfer with the environment.

Thermal cycles of the disclosure (e.g., the cycles in FIGS. 4 and 5 ) may be implemented through various configurations of pipes and valves for transporting the working fluid between the turbomachinery and the heat exchangers. In some implementations, a valving system may be used such that the different cycles of the system can be interchanged while maintaining the same or nearly the same temperature profile across at least one, across a subset or across all of counter-flow heat exchangers in the system. For example, the valving may be configured such that the working fluid can pass through the heat exchangers in opposite flow directions on charge and discharge and flow or conveyance directions of the HTS and CTS media are reversed by reversing the direction of the pumps or conveyance systems.

In some implementations, the system may be set up to enable switching between different cycles. Such a configuration may be advantageous as it may reuse at least a portion, or a substantial portion, or a majority, of the same piping and/or connections for the working fluid in both the charging and discharging modes. While the working fluid may change direction between charge and discharge, the temperature profile of the heat exchangers can be kept constant, partially constant, or substantially or fully constant, by changing the direction in which the HTS medium and the CTS medium are pumped or conveyed when switching from charge to discharge and vice-versa, and/or by matching the heat fluxes of the working fluid, the HTS medium and the CTS medium appropriately.

III. Illustrative Baffled Thermoclines in a Brayton Cycle Engine

FIG. 7 illustrates a Brayton cycle heat engine configured to generate electrical power and supply such power to an electrical grid. The heat engine may be reversible (i.e., operate as a heat pump) and may take the form of other heat engines and/or reversible heat engines describe herein and may include additional or alternative components than those shown in the illustration.

The heat engine may include a generator/motor 701 that may generate electricity or use electricity to operate a compressor 703. The generator/motor 701 may be mechanically coupled to the compressor 703 and a turbine 705. The compressor 703 and the turbine 705 may be coupled to the generator/motor 701 via one or more shafts 715. Alternatively, the compressor 703 and the turbine 705 may be coupled to the generator/motor 701 via one or more gearboxes and/or shafts. The heat engine may use mechanical work to store heat and/or may provide mechanical work from stored heat. The heat engine may have a hot side 717 and a cold side 719.

In one embodiment, the heat engine may include a hot-side thermocline vessel 707 coupled between the compressor 703 and the turbine 705 on the hot side 717. The hot-side thermocline vessel 707 may act as a direct-contact heat exchanger, where a working fluid is in direct contact with a solid thermal medium and at greater than atmospheric pressure. An optional recuperative heat exchanger 711 may be disposed in the working fluid path between the compressor 703 and the hot-side thermocline vessel 707. With the use of the solid thermal medium, which may be effective across a wide temperature range, it may be possible to reduce or eliminate the use of a recuperative heat exchanger.

A cold-side thermocline vessel 709 may be coupled between the turbine 705 and the compressor 703 on the cold side 719. The cold-side thermocline vessel 709 may act as a direct-contact heat exchanger, where a working fluid is in direct contact with a solid thermal medium and at less than atmospheric pressure, wherein the solid thermal medium on the cold side may be different than the solid thermal medium on the hot side. The recuperative heat exchanger 711 may be disposed in the working fluid path between the turbine 705 and the cold-side thermocline vessel 709, such that a working fluid stream downstream of the turbine 705 is in thermal contact with a working fluid stream downstream of the compressor 703.

The hot-side thermocline vessel 707 and the cold-side thermocline vessel 709 are preferably insulated pressure vessels. As used herein, a pressure vessel is intended to refer to a vessel or containment area that can operate at either or both above atmospheric pressure (e.g., 1 to 5 bar, 5 to 30 bar, 30 to 100 bar, or greater) and/or below atmospheric pressure (e.g., 1×10⁵ to 3×10³ Pa, 3×10³ to 1×10⁻¹ Pa, 1×10⁻¹ to 1×10⁻⁷ Pa, or less). They may be insulated to prevent or reduce transmission of heat contained within the vessel to the external environment. They may further be sealed to maintain the pressure of incoming working fluid that may be substantially above or below atmospheric pressure and to maintain a substantially isobaric environment where the working fluid may directly contact the solid thermal medium. The thermocline vessels 707 and 709 may include one or more inlets for receiving the working fluid at non-atmospheric pressure from the closed thermodynamic cycle system, such as a Brayton cycle system, and one or more outlets for dispatching the working fluid at non-atmospheric pressure to the the closed thermodynamic cycle system. The inlets and outlets may be one or more apertures through the exterior walls of the thermocline vessels 707 and 709 and that are connected to the respective working fluid streams and sealed from the atmosphere.

The thermocline vessels 707 and 709 each preferably contain a solid thermal medium. The solid thermal medium may have a structure with porosity sufficient to allow the working fluid to flow through the solid thermal medium. The solid thermal medium may be segregated into a plurality of zones that are defined by a plurality of baffle structures within each pressure vessel. The baffle structures may be configured to limit direct transfer of heat between the solid thermal medium in different zones, whether by conductive or radiative means. The baffles may have one or more fluid channels that allow the working fluid to flow past the baffles and in contact (direct or thermal) with the solid thermal medium. The baffles may allow the thermocline vessels 707 and 709 to maintain a thermocline within the vessel for an extended period of time by reducing the transfer of heat from a hot side of the thermocline to the cold side of the thermocline. As an illustrative example, the thermocline in the hot-side thermocline vessel 707 may exhibit a temperature difference of approximately 1500° C., 1400° C., 1300° C., 1200° C., 1100° C., 1000° C., 900° C., 800° C., 700° C., 600° C., 500° C., 400° C., 300° C., or 200° C. For example, the temperature (T_(h_high)) of the solid thermal medium near the outlet may be approximately 1500° C., 1400° C., 1300° C., 1200° C., 1100° C., 1000° C., 900° C., 800° C., 700° C., 600° C. and the temperature (T_(h_low)) of the solid thermal medium near the inlet may be approximately 200° C. or 100° C. As another example, the thermocline in the cold-side thermocline vessel 709 may exhibit a temperature difference of approximately 400° C., 300° C., 200° C., or 100° C. For example, the temperature (T_(c_high)) of the solid thermal medium near the inlet may be approximately 200° C., 100° C., 70° C., 30° C. or 0° C. and the temperature (T_(c_low)) of the solid thermal medium near the outlet may be approximately −30° C., −100° C., or −200° C. Each of the thermocline vessels may have one or more pressure sealed access ports to load or unload the solid thermal medium for thermal charging, maintenance, or other access requirements.

The heat engine illustrated in FIG. 7 may also have fluid paths configured to allow it to operate without a recuperator (as in FIG. 2 ) and/or to operate reversibly and function to store excess electrical energy in the form of thermal energy, similar to the cycle shown in FIG. 4 or FIG. 1 (without a recuperator), where the hot side heat exchanger 2 and associated tanks 6 and 7 and HTS medium 21 are replaced with thermocline 707 and the cold side heat exchanger 4 and associated tanks 8 and 9 and CTS medium 22 are replaced with thermocline 709, and the fluid flow paths are as indicated in FIG. 1, 2 or 4 . Due to inefficiencies likely present in the system, excess heat may need to be rejected in the discharge or charge cycles. Heat rejection devices may be inserted into the fluid paths of the described embodiments without departing from the claimed subject matter.

As an example embodiment only, in a discharge cycle, a heat rejection device 713, such as a cooling tower, may be disposed in, or coupled to, the working fluid stream between the turbine 705 and the cold-side thermocline vessel 709. The heat rejection device 713 may eject heat from the system, where the heat may be carried into the heat rejection device 713 by the working fluid and ejected to the atmosphere or other heat sink.

FIG. 8 illustrates a schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment. The thermocline vessel 800 may include a pressure vessel 806 that is insulated. The pressure vessel 806 may take various forms sufficient to withstand the pressure of the working fluid and to prevent or reduce heat transfer between the solid thermal medium and the external environment. For example, the pressure vessel 806 may be a container. The outside walls of the container may include one or more materials designed to withstand pressure and/or to minimize heat transfer. For example, the walls may include, internal insulation, an interior surface of refractory material, a structural steel core, and an external insulation and/or protective material capable of withstanding long-term environmental exposure. Pressure sealed access ports may be included within the walls.

The thermocline vessel 800 may include an inlet 802 for working fluid from the Brayton cycle system and an outlet 804 for working fluid to the Brayton cycle system. The inlet 802 and outlet 804 may each be simple pipe ports with an opening into the interior of the pressure vessel 806 and/or they may include more complex structures such as distribution plenums that connect to external piping containing the working fluid.

The thermocline vessel 800 may include baffle structures in one or more configurations, such as baffle structures 808, 810, and 812. Each of the baffle structures 808, 810, 812 may be an insulating wall and may include fluid channels 814 that allow passage of working fluid past the baffle structures. For clarity of the illustration in FIG. 8 , only a few of the fluid channels 814 are labeled. The fluid channels 814 may be apertures through the baffle structure 810. As illustrated in FIG. 8 , the baffles 808, 810, 812 may extend completely across the thermocline vessel 800. Baffle structure 814 may partially or completely define an inlet region 816 that may be filled with incoming working fluid from the inlet 802. Baffle structure 810 may include an arrangement of fluid channels 814 that allows widespread distribution of the working fluid across a span of the interior of the thermocline vessel 800. For example, baffle structure 810 may have more or less fluid channels 814 or differently spaced or configured fluid channels 814 than other baffle structures in the thermocline vessel 800. Similarly, baffle structure 812 may partially or completely define an outlet region 818 that may be filled with incoming working fluid. Baffle structure 812 may include an arrangement of fluid channels 814 that allows widespread collection of the working fluid from across a span of the interior of the thermocline vessel 800. For example, baffle structure 812 may have more or less fluid channels or differently spaced or configured fluid channels than other baffle structures in the thermocline vessel 800.

Baffle structures 808, together and/or in conjunction with baffle structures 810, 812 may partially or completely define a plurality of zones within the insulated pressure vessel 800. Each zone may contain a solid thermal medium 820. The baffle structures 808 are preferably configured to limit direct transfer of heat between the solid thermal medium 820 in different zones, such as conduction or radiation of heat between a bolus of the solid thermal medium in one zone and a bolus of the solid thermal medium in a different zone, while allowing passage of a working fluid past the baffle structures and into contact with the solid thermal medium 820.

In the thermocline vessel 800, each of the zones may contain the solid thermal medium 820 at a different temperature, forming a thermocline configuration as illustrated in FIG. 8 by T₂ through T₈. A zone proximate to the inlet (i.e., an “inlet zone”) may be at a temperature T₂. A zone proximate to the outlet (i.e., an “outlet zone”) may be at a different temperature T₉. Each of the zones between the inlet zone and the outlet zone may be at differing temperatures that form a gradient between T₂ and T₈. For example, in a hot-side thermocline vessel, the thermocline may take the form of T₉>T₈>T₇>T₆>T₅>T₄>T₃>T₂. As further illustration, the solid thermal medium 820 in the outlet zone may be at approximately T₉=600° C. and the solid thermal medium 820 in the inlet zone may be at approximately T₂=400° C. As another example, in a cold-side thermocline vessel, the thermocline may take the form of T₉<T₈<T₇<T₆<T₅<T₄<T₃<T₂. As further illustration, the solid thermal medium 820 in the outlet zone may be at approximately T₉=−70° C. and the solid thermal medium 820 in the inlet zone may be at approximately T₂=30° C.

FIG. 9 illustrates another schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment. The thermocline vessel 900 illustrated in FIG. 9 is similar to the embodiment in FIG. 8 except that the baffle structures 908 do not extend completely across thermocline vessel 800, as illustrated in FIG. 8 . The baffle structures 908 may be insulating walls and the fluid channels 914 may be formed by spaces between the ends of the insulating walls and an interior wall of the pressure vessel. The resulting serpentine flow path for the working fluid may provide a greater residence time for the working fluid in the thermocline vessel 900 when compared to the more direct flow path of the thermocline vessel 800.

FIG. 10 illustrates another schematic arrangement, in cut-away view, of a baffled thermocline pressure vessel according to an example embodiment. The thermocline vessel 1000 is similar to the embodiments in FIGS. 8 and 9 except that the baffle structures 1008 are insulated chambers which define zones for the solid thermal medium 820 within the chambers. Each of the baffle structures 1008 include one or more fluid channels 1014 that allow a working fluid to enter the chambers and contact the solid thermal medium 820. In one example, the baffle structures 1008 may be formed as permanent constructions, for examples with walls of refractory brick or other materials with gaps or apertures through the walls. In another example, the baffle structures 1008 may be mobile containers such as crucibles that may be placed into, removed from, or moved within the pressure vessel 806. Such an arrangement may reduce the overall heat transfer between the working fluid and the solid thermal medium 820, but may have other benefits such as flexibility of design, ease of maintenance, and/or recharging the thermal energy in the system.

FIG. 11 illustrates another schematic arrangement, in cut-away view, of a baffled thermocline vessel with a runner system according to an example embodiment. The thermocline vessel 1100 illustrated in FIG. 11 is similar to the embodiment in FIG. 8 except that the working fluid is confined within runners instead of flowing throughout the vessel 1106. Accordingly, vessel 1106 may be operable at atmospheric pressure. The working fluid may enter the vessel 1100 at inlet 1102, be distributed through a plenum 1116 to a series of runners 1122, enter a plenum 118, and exit the vessel 1100 at outlet 1104. The runners 1122 are configured to allow heat to transfer between the solid thermal medium 820 and the working fluid in the runners 1122. Heat transfer may be by conduction, convection (e.g., from an atmosphere within the vessel 1106), and/or radiative action. Preferably the runners 1122 are in direct contact with the solid thermal medium, but they may be in non-direct thermal contact only.

As an example only, the baffle structures 1108 may be insulating walls similar to those described with with respect to FIGS. 8 and 9 , except that the fluid channels 1114 are sized to allow the runners containing the working fluid to pass the walls.

FIGS. 12 through 15 illustrate example embodiments of forms of baffle structures inside a thermocline vessel, as viewed in a direction of working fluid flow. Each of the illustrations reflects implementation in a pressure vessel 806, but the example baffle structures could be implemented in a non-pressurized thermocline vessel such as the runner system illustrated in FIG. 11 . FIG. 12 illustrates a baffle structure 1200 a where the fluid channels are apertures 1202 a between components 1202 (e.g., bricks) used to construct the baffle structure. FIG. 13 illustrates a baffle structure 1200 b where the fluid channels are apertures 1204 a through a perforated material 1204. FIG. 14 illustrates a baffle structure 1200 c where the fluid channels are gaps 1206 a between components 1206 (e.g., slats or rods) used to construct the baffle structure. FIG. 15 illustrates a baffle structure 1200 d where the fluid channel is a gap 1208 a between an interior surface of the pressure vessel 806 and the wall 1208.

Preferably the baffle structures 1200 a, 1200 b, 1200 c, and 1200 d have insulating properties to limit the conduction of heat from the solid thermal medium on one side of the structure to the solid thermal medium on the other side of the structure. The fluid channels may be sized or located to limit or prevent the movement of the solid thermal medium through the aperture. Alternatively or additionally, the fluid channels may be sized, located, and/or shaped to limit the passage of radiative heat transfer. For example, a fluid channel may be smaller than a pellet of the solid thermal medium and/or a fluid channel may be angled or convoluted in form such that there is no or limited line-of-sight between the solid thermal medium on one side of the baffle structure and the solid thermal medium on the other side of the baffle structure.

VI. Conclusion

While various aspects and embodiments have been disclosed herein, other aspects and embodiments will be apparent to those skilled in the art. The various aspects and embodiments disclosed herein are for purposes of illustration and are not intended to be limiting, with the true scope and spirit being indicated by the following claims. 

What is claimed is:
 1. An energy system, comprising: a turbomachinery system comprising a compressor configured to compress working fluid, and a turbine configured to expand the working fluid to generate mechanical rotation, wherein in a charge mode, the compressor receives rotational drive from a motor to compress the working fluid to store thermal energy, and in a discharge mode the turbine transmits at least a portion of mechanical rotation to drive a generator to generate electrical energy; a hot-side thermocline vessel comprising a first plurality of zones of a first solid thermal medium defined by first baffle structures in an interior of the hot-side thermocline vessel, wherein the first baffle structures are configured to limit direct transfer of heat between the first solid thermal medium in different zones, wherein the first solid thermal medium within a first inlet zone proximate to an inlet of the working fluid to the hot-side thermocline vessel is at a first temperature, and wherein the first solid thermal medium within a first outlet zone proximate to an outlet of the working fluid from the hot-side thermocline vessel is at a second temperature different than the first temperature; and a cold-side thermocline vessel comprising a second plurality of zones of a second solid thermal medium defined by second baffle structures in an interior of the cold-side thermocline vessel, wherein the second baffle structures are configured to limit direct transfer of heat between the second solid thermal medium in different zones, wherein the second solid thermal medium within a second inlet zone proximate to an inlet of the working fluid to the cold-side thermocline vessel is at a third temperature, and wherein the second solid thermal medium within a second outlet zone proximate to an outlet of the working fluid from the second thermocline vessel is at a fourth temperature different than the third temperature, and a working fluid cycle path arranged, in the discharge mode, for circulating the working fluid in a closed cycle through, in sequence, the compressor, the hot-side thermocline vessel in thermal contact with the first solid thermal medium, the turbine, and the cold-side thermocline vessel in thermal contact with the second solid thermal medium.
 2. The energy system of claim 1, wherein the first plurality of zones comprise one or more intermediate zones disposed between the first inlet zone and the first outlet zone.
 3. The energy system of claim 2, wherein the first solid thermal medium in each of the one or more intermediate zones is at a temperature between the first temperature and second temperature.
 4. The energy system of claim 1, wherein the magnitude of the difference between the first temperature and the second temperature is greater than 200° C.
 5. The energy system of claim 1, further comprising a heat rejection device configured to eject from the energy system heat carried by the working fluid.
 6. The energy system of claim 1, wherein the working fluid within the hot-side thermocline vessel is at a pressure greater than atmospheric pressure.
 7. The energy system of claim 1, wherein the working fluid within the cold-side thermocline vessel is at a pressure greater than atmospheric pressure.
 8. The energy system of claim 1 further comprising a recuperative heat exchanger configured to thermally contact the working fluid downstream from the compressor with the working fluid downstream from the turbine, wherein in the discharge mode the working fluid circulates, in order, through (i) the compressor, (ii) the recuperative heat exchanger, (iii) the hot-side thermocline vessel in thermal contact with the first solid thermal medium, (iv) the turbine, (v) the recuperative heat exchanger, and (vi) the cold-side thermocline vessel in thermal contact with the second solid thermal medium.
 9. The energy system of claim 8 further comprising a heat rejection device configured to eject from the system heat carried by the working fluid, wherein in the discharge mode the working fluid circulates, in order, through (i) the compressor, (ii) the recuperative heat exchanger, (iii) the hot-side thermocline vessel in thermal contact with the first solid thermal medium, (iv) the turbine, (v) the recuperative heat exchanger, (vi) the heat rejection device, and (vii) the cold-side thermocline vessel and the second solid thermal medium. 